Pumping crude oil by centrifugal impeller having different blade … · 2017-04-10 · Pumping...
Transcript of Pumping crude oil by centrifugal impeller having different blade … · 2017-04-10 · Pumping...
ORIGINAL PAPER - PRODUCTION ENGINEERING
Pumping crude oil by centrifugal impeller having different bladeangles and surface roughness
Sayed Ahmed Imran Bellary1 • Abdus Samad1
Received: 19 January 2015 / Accepted: 17 May 2015 / Published online: 9 June 2015
� The Author(s) 2015. This article is published with open access at Springerlink.com
Abstract Centrifugal pumps are extensively used in the
oil and gas industries and the pump performance drops
with higher viscosity and higher surface roughness of the
pump impeller, and the impeller design parameters have
significant effect on the pump performance. Through the
present research, crude oil pumping behavior has been
predicted, analyzed and compared with other fluids. A 3D
flow simulation using Reynolds-averaged Navier–Stokes
(RANS) equation was performed by considering different
blade angles and impeller surface roughness to pump crude
oil, kerosene, gasoline, saline-water and water. Standard k-
e two-equation turbulence model was used for the turbulent
closure of steady incompressible flow. The investigation
shows that the blade angles have significant influence on
the head, input power and efficiency of the impeller for
different liquids. Higher head and power, and lower
hydraulic efficiency were observed with higher surface
roughness values.
Keywords Centrifugal impeller � Exit blade angle �Viscosity � Roughness � Slip factor
List of symbols
Abbreviations
RANS Reynolds-averaged Navier–Stokes
CFD Computational fluid dynamics
PS Pressure side
SS Suction side
LE Leading edge
TE Trailing edge
Description
B Blade width (mm)
C Absolute fluid flow velocity (m/s)
ceq Equivalence factor
cm Meridional velocity component (m/s)
cu Peripheral velocity component (m/s)
D Diameter (m)
f Fluid
fi Body force (N)
fR Roughness effect
g Acceleration due to gravity (m/s2)
H Head generated (m)
DH Hydraulic losses (m)
ks Sand roughness (lm)
ks? Roughness Reynolds number
m Mass flow rate (kg/s)
N Impeller speed (rpm)
P Power consumed by pump (kW)
p Pressure (N/m2)
Q Volume flow rate (m3/s)
Re Reynolds number
r Radius (m)
t Total blade thickness (mm)
Uj Three-dimensional velocity vector
U Peripheral velocity (m/s)
us Frictional velocity (m/s)
& Abdus Samad
Sayed Ahmed Imran Bellary
1 Wave Energy and Fluids Engineering Laboratory,
Department of Ocean Engineering, Indian Institute of
Technology Madras, Chennai 600036, India
123
J Petrol Explor Prod Technol (2016) 6:117–127
DOI 10.1007/s13202-015-0173-y
w Relative fluid velocity (m/s)
wu Peripheral component of w (m/s)
ws Relative velocity on suction side (m/s)
y Blade thickness
z Blade number
Greek symbol
b Blade angle (o)
e Rate of kinetic energy dissipation (J/s)
g Hydraulic efficiency (%)
k Turbulence kinetic energy (J)
l Dynamic viscosity (N s/m2)
m Kinematic viscosity (m2/s)
q Density of fluid (kg/m3)
r Slip factor
sij Viscous stress tensor
sw Shear stress at the wall (N/m2)
Subscript
1 Inlet
2 Outlet
a Actual
h Hub
o Eye
s Shaft
th Theoretical
max Maximum
Introduction
Now-a-days centrifugal pump either as single stage or
multistage is extensively used in upstream, midstream and
downstream oil industries. For example, the upstream oil
industry uses to lift fluid from the wellbore, to deliver fluid
in the separation system, etc. Performance characteristics
of a pump greatly depend on geometry and surface prop-
erty of an impeller. Again, the viscosity which can be
defined as resistance to flow has significant impact on head,
efficiency and power consumption of the pump.
Many researchers reported analysis of centrifugal pump
for the flow behavior, influence of geometric parameters,
etc. (Kamimoto and Matsuoka 1956; Varley 1961; Ste-
panoff 1940; Lazarkiewicz and Troskolanski 1965). The
works on inlet and exit blade angle shows that the per-
formance can be altered, when the angles are modified. A
higher exit blade angle was suggested by the researchers.
Kamimoto and Matsuoka (1956) experimentally investi-
gated the effect of exit blade angles and reported that the
impeller with 30� exit angle has the best performance. For
a double suction centrifugal pump, the head increment can
be achieved by increasing exit blade angle and an
improvement in efficiency by varying exit blade angles can
be obtained (Varley 1961). The inlet blade angle modifi-
cation of a radial impeller has been reported by Sanda and
Daniela (2012), and Luo et al. (2008). They reported that
performance enhancement is possible by this modification.
Normally, pump designers design the pumps for water
and they rarely test for different viscosities. Obviously,
higher viscosity has detrimental effect on performance as
the higher viscosity has more resistance to flow (Murakami
et al. 1980; Telow 1942; Ippen 1946; Itaya and Nishikawa
1960). Again, the density effect has not been considered by
the researchers. If the fluid is having higher viscosity or
density it is inevitable to check amount of extra power
consumption required and this will help energy auditing of
a company. The power requirement of pump selection on
the basis of the head requirement can be evaluated. Reports
on centrifugal pump handling viscosity liquids shows that
that the large exit angle exhibits an improvement in head
and efficiency (Aoki et al. 1947; Ohta and Aoki 1996; Fard
and Boyaghchi 2007). Li (2011) investigated the effect of
various blade angles of an industrial oil and concluded that
the blade exit angle has equal effects on head, power and
efficiency. Shojaeefard et al. (2012) investigated both
experimentally and numerically the effect of impeller exit
angle for oil as working fluid and showed that increase of
impeller outlet angle leads to improvement in the perfor-
mance. Shigemitsu et al. (2011) performed experiment and
numerical analysis on a mini turbo-pump and reported that
the flow rate of the maximum efficiency shifts to a large
flow rate due to the increase of the blade exit angle.
If any centrifugal pump works for a longer period, the
surface gets deteriorated because of fouling, cavitation or
erosion. The oil industry typically handles multi-phase flow
and the phases are oil, water gas and sand. There may be
some corrosive gasses also such as H2S. The acidic gas H2S
reacts with water and forms H2SO4 which is highly reactive
to the metal. Similarly CO2 or Cl gives ion for corrosion of
metal. These corrosive gasses along with sand-jetting effect
accelerate the erosion. A small amount of sand can initiate
pitting on the surface and the new surface is attacked by the
reactive gasses or acids. Hence, the pump handling fluid
from the well bore or at the surface production operations
or at the downstream petroleum processing industries will
develop micro-pitting and as a result there will be surface
roughness. On one side the oil or the liquid hydrocarbon
(HC) coats the surface and helps reducing the reaction, but
at the same time it resists the flow through the pump. The
high-density fluid can have higher head but the pump may
be running under off-design condition. Hence, finally pump
performance will be dropped.
Experiments for different surface roughness values were
conducted by Varley (1961), but they did not report the
inter-dependency of exit blade angle with surface
118 J Petrol Explor Prod Technol (2016) 6:117–127
123
roughness. Fard and Boyaghchi (2007) studied the influ-
ence of various blade exit angles to handle viscous fluids
by computational and experimental methods. The investi-
gation did not include roughness of wet wall. Li (2011)
reported effect of exit blade angle, viscosity and roughness
by CFD simulations, but the effect of inlet and outlet blade
angle on roughness was not included.
Several researchers (Johnston and Rothe 1967; Johnston
et al. 1972; Bayly and Orszag 1988) reported a low-energy
separated region, i.e., a boundary layer on pressure side
which is unstable and tends to propagate from the hub and
shroud surfaces on the suction side in the impeller. This
kind of flow pattern develops jet wake flow. The analytical
and experimental results (Dean and Senoo 1960; Tuzson
1993) show that a separated region because of acceleration
and corresponding pressure increase from suction side to
pressure side forming a separated region has its limitations.
This kind of jet wake flow pattern is exhibited only under
certain conditions.
Thus, the purpose of the present study is to investigate
the combined effect of blade angles and surface roughness
on impeller performance by numerical simulation for crude
oil, kerosene, gasoline, saline-water, and water at different
flow rates. The flow mechanism, nature and distribution of
velocity and pressure in the pump at design and off-design
point were reported. Reason to find performance change
due to viscosity effect in a centrifugal pump was performed
at different conditions.
Numerical formulation
In the present problem, a centrifugal pump impeller was
considered (Fig. 1; Table 1). The detail geometry of the
impeller corresponds to the values obtained from the con-
ventional design procedure of impeller design as given in
refs. (Lazarkiewicz and Troskolanski 1965; Stepanoff
1964; Church 1972). The same geometry has already been
reported by the authors (Bellary et al. 2014; Bellary and
Samad 2014). Design procedures (Lazarkiewicz and
Troskolanski 1965; Church 1972) and the input values such
as head (H), discharge (Q), and impeller speed (N) were
considered for the same reference. The blades had constant
thickness (y) and width (b) with trailing edge chamfered.
To investigate the influence of blade angles on the
performance of centrifugal impeller, three inlet blade
angles (b1 = 17�, 23� and 28�) and three outlet blade
angles (b2 = 25�, 40� and 70�) were considered. Surface
roughness was changed from hydrodynamically smooth
surface to a sand roughness of 50 and 100 lm. Here,
hydrodynamically smooth surface refers to a roughness of
5 lm. Table 2 represents the liquids used for the analyses.
All the simulations were performed at 20 �C temperature.
X
u1Wu1
W1
β1Y
cm1= c1
α1
Z
B
A
D Cβ2 α2
cm2W2
u2
Wu2
c2
cu2
(a) Inlet and exit velocity triangle
(b) Meridional configuration for impeller
b1
rh
ro
r2
r1
b2
Trailing edge
Leading edge
Hub
Inlet
Blade
(c) Impeller
Periodic
Blade
Shroud
Outlet
Hub
Inlet
(d) Flow domain and mesh
Fig. 1 Computational domain
J Petrol Explor Prod Technol (2016) 6:117–127 119
123
CAD modeling and meshing of the flow domain were
carried out by using Ansys-BladeGen and -TurboGrid
module, respectively (Ansys-CFX 13.0 2010). Ansys-CFX
13.0 (2010) was used for the flow simulations. The total
number of nodes, elements, convergence criteria and iter-
ation steps, etc., are given in Table 3. The mesh in the flow
passage was with O-grid block, and it was with H-grid
block at the inlet and outlet. To achieve linear velocity
distribution in the viscous sublayer, higher number of cells
was employed near the wall. To avoid fine mesh incon-
sistencies and to resolve the boundary layer, at least 10
nodes were included into the boundary layer (Ansys-CFX
13.0 2010; Gulich 2010). In all simulations performed the
size of elements used next to wall was 11.13. Figure 2
shows head variation with the number of nodes of the flow
domain.
The Reynolds-averaged Navier–Stokes (RANS) equa-
tions for steady, incompressible turbulent flow of constant
properties fluids can be written in tensor form as:
oUi
oxi¼ 0 ð1Þ
qf Uj
oUi
oxj¼ � oP
oxiþ o
oxjðlf þ ltÞ
o2Ui
oxjoxj
� �� �ð2Þ
Here, turbulent viscosity is modeled by the k-e model
(turbulence model) as:
lt ¼ Clqfk2
eð3Þ
The equations for turbulent kinetic energy and
dissipation rate can be written as:
qf Uj
ok
oxj¼ o
oxjlf þ
ltrk
� �ok
oxj
� �þ Pk � qf e ð4Þ
qf Uj
oeoxj
¼ o
oxjlf þ
ltre
� �oeoxj
� �þ ekðCe1Pk � Ce2qf eÞ;
ð5Þ
where
Pk ¼ ltoUi
oxjþ oUj
oxi
� �oUi
oxjð6Þ
The values of the closing constants are Cl = 0.09,
Ce1=1.44, Ce2 = 1.92, rk = 1.0, and re = 1.3.
To account for Reynolds number (Re[ 105), the stan-
dard k-e turbulence model was employed. Within Ansys-
CFX (2010), the k-e turbulence model uses the scalable
wall-function approach to improve robustness and accu-
racy. High-resolution upwind discretization scheme was
Table 1 Features of impeller
Parameter Dimension
Shaft diameter, Ds 40 mm
Eye diameter, Do 182 mm
Hub diameter, Dh 55 mm
Inlet diameter, D1 160 mm
Inlet blade width, b1 54 mm
Outlet blade width, b2 30 mm
Inlet blade angle, b1 17�, 23� and 28�Outlet blade angle, b2 25�, 40� and 70�Blade number, z 7
Outlet diameter, D2 365 mm
Table 2 Viscosity and density of different fluids
Fluid Viscosity (N s/m2) Density (kg/m3)
Water 1.002E-3 997
Saline-water 1.080E-3 1031
Crude oil 5.000E-3 835
Gasoline 5.000E-4 720
Kerosene 2.100E-3 810
Table 3 Meshing and boundary conditions
Parameter Description
Flow domain Single impeller
Interface Periodic
Mesh/nature Structural/hexahedral
Nodes 634,161
Elements 568,620
Fluid nature Water, saline-water, crude oil, gasoline and
kerosene
Inlet Pressure
Outlet Mass flow rate
Residual convergence
value
1 9 10-5
Time taken for
simulation
15 h
Iteration steps 2000
Mass imbalance 0.0001 %
Fig. 2 Head variation with number of nodes at design flow rate
(Q = 0.1003 m3/s)
120 J Petrol Explor Prod Technol (2016) 6:117–127
123
used for solving convection terms with central difference
schemes for diffusion terms. Ansys-CFX (2010) solver is a
coupled solver and solves the hydrodynamic equations for
velocity and pressure (u, v, w, p) as a single system. The
simulations were carried out on 3.4 GHz core i7-3370
processor with 8 GB ram. Time per iteration was within
21 s.
Backward swept (b2\ 90o) blades are used in pumps
and blowers (Srinivasan 2008). The blade angles have
influence on the total head and efficiency of pump. As per
Euler’s equation, head generated by an impeller blade is
given by
H ¼ cu2u2 � cu1u1
g: ð7Þ
Normal entry at impeller inlet implies cu1 = 0 and from
velocity triangle shown in Fig. 1a cu2 can be given as:
cu2 ¼ u2 �Q cot b2pDb2
: ð8Þ
From Fig. 1a it is clear that as b2 increases, the absoluteexit velocity c2 also increases. Increase in c2 causes cu2 to
increase which results in increase in head.
Slip factor (r) generally explains the slip effect at the
impeller exit and the factor is used to specify theoretical
head developed by an impeller (Srinivasan 2008). The slip
factor is estimated by the relation,
r ¼ 1� p sin b2z
: ð9Þ
Input power in terms of cu2 is given by
P ¼ qgQu2cu2g
: ð10Þ
The Eqs. (8) and (10) indicate that the increase in exit
angle results in increase in cu2 which increases the input
power.
Hydraulic efficiency is given by
g ¼ Ha
Hth
¼ Hth � DHHth
: ð11Þ
The hydraulic losses and theoretical head (Hth) at design
point differ from the off-design conditions. The head and
the corresponding efficiency are higher for a larger exit
angle (Srinivasan 2008).
Surface roughness
Surface roughness increases the flow resistance in turbulent
flow. It has no effect on the resistance due to no exchange
of momentum across the flow in laminar flow condition. A
wall is termed as hydraulically smooth if all the roughness
peaks are within the laminar sublayer. Increase in Reynolds
number (Re) results in decrease in boundary layer thick-
ness. A hydraulically rough surface has some peaks greater
than the boundary layer thickness. Surface roughness is
expressed in term sand roughness (ks). The sand roughness
considers a uniform surface structure while the surfaces
may have irregular roughness which considers maximum
roughness depth ksmax (Gulich 2010; Tuzson 2000) in
actual scenario. The relationship between ks and ksmax is
given by
ceq ¼ksmax
ks: ð12Þ
As per Tuszon (2000), the roughness Reynolds number
ks? can be defined as the ratio of frictional force to the
viscous force, and given by
kþs ¼ usks
mð13Þ
us ¼ffiffiffiffiffiswq
r: ð14Þ
Disc friction losses become significant for higher ks and
the hydraulic losses as well as efficiency get increased
(Eq. 11). Energy losses in impeller depend on Reynolds
number and relative surface roughness and termed as
roughness effect (Gulich 2010). The roughness effect (fR)
on disc friction can be given as:
fR ¼log 12:5
Re
log 0:2 ksr2þ 12:5
Re
� �8<:
9=;
2:15
: ð15Þ
Results and discussion
Problem setup
Initially, the pump geometry was designed and grid-de-
pendency test (Fig. 2), check for turbulence model and
validity of the CFD simulations were checked for water at
design flow condition. The number of nodes was 550,000
for all the simulations as the variation of head with the
increase in number of nodes were not significant. Further it
was checked for accuracy of the CFD model and found that
the results were predicting well (Bellary et al. 2014; Bel-
lary and Samad 2014). After that the geometry was chan-
ged for different inlet and exit angles, and different surface
roughness. The simulations were done for different fluids
and for a wider operating range.
Effect of blade angles
The simulations were carried out at different mass flow rates
and initially parametric study was performed for different
J Petrol Explor Prod Technol (2016) 6:117–127 121
123
inlet and exit angles (b1 = 17�–28� and b2 = 25�–70�).Figure 4 represents the performance curves for different
combinations of the inlet and exit angles. The impeller with
b1 = 23� has slightly better performance as compared to
b1 = 17� and 28o designs, hence all further calculations
were performed at b1 = 23� with different b2. Figure 5
demonstrates the head, shaft power and hydraulic efficiency
at different flow rates. Large b2 results in an increase in
head. This is due to the increase in absolute flow velocity at
exit (c2) and the corresponding peripheral velocity (cu2), i.e.,
dynamic part of the head increases more rapidly.
From Eq. (7) it is obvious that the head increases with
the increase in cu2. Equation (8) represents a drooping
down straight line and says that a larger discharge angle
produces a higher head than a smaller exit angle. The
maximum change of the head curve due to the variation in
discharge angle is almost same for the fluids. This implies
that the effect of discharge angle on the head curve is
independent of the fluid viscosity (l = 5.0E - 4 to
50.0E - 4 N-s/m2). Irrespective of viscosity, the maxi-
mum change in head due to the change in b2 remains
almost same (Fig. 5). Thus, effect of discharge angle on
head is independent of the viscosity of fluid being used.
For the same flow rate and discharge angle, the input
power is more for denser fluids (Fig. 5b) and this is
explained by Eq. (10). The increase in flow rate results in
increase in power consumption. The density of crude oil,
gasoline or kerosene is lesser than that of water or saline-
water, which causes the lesser power consumption. Higher
b2 increases cu2; hence, the power consumption by the
impeller is more (Eq. 10).
A large exit angle favors efficiency enhancement
(Fig. 5). The efficiency is maximum at the best efficiency
point and gets reduced uniformly to the right of the best
efficiency point. At this point, secondary and profile losses
are minimum and at the off-design points the secondary
and shock losses increases which increases total hydraulic
losses resulting in reduced efficiency (Srinivasan 2008).
The efficiency for crude oil, saline-water, gasoline or ker-
osene is lower than that for water. The decrease in effi-
ciency, while pumping the crude oil and saline-water is due
to the disc friction losses over the outsides of the impeller
shroud and hub due to viscous effect. The results agree
with the results of Gulich (2010) and Li (2008).
Contours of static pressure and liquid flow velocity for
design point (Q = 0.1003 m3/s) at 50 % span are shown in
Figs. 6 and 7. The static difference between impeller inlet
and outlet increases with large b2. This happens because offluid flow velocity at impeller outlet, which decreases with
increase in discharge angle. The contours depict a smooth
flow and the pressure increases continuously towards the
exit of the domain. The lowest static pressure can be
observed at the impeller inlet on suction side (Fig. 6). At this
location, usually an inception of cavitation appears. In the
present work, the fluids were at the above saturation pressure
and the occurrence of cavitation is ruled out (Gulich 2010).
The static pressure and kinetic energy reaches a peak
value at the impeller exit. This is due to the energy transfer
by the impeller to the fluid. Lowest pressure exists at the
suction side (near the leading edge) of the impeller. With
the increase in blade angle, the pressure difference between
outlet and inlet increases. The liquids with large b2 have
higher pressure difference because of the decrease in
velocity at the impeller outlet for large b2 (Fig. 5). This
agrees with the existing analytical results (Lazarkiewicz
and Troskolanski 1965; Srinivasan 2008).
For a radial blade impeller jet wake flow pattern can be
observed if sin(b) is relatively lesser than the ratio of the
average velocity on the seperating streamline and circum-
ferential velocity (Tuzson 2000). However, with backward
curved vanes, the blades lean strongly backwards making
sin(b) relatively large and the above statement does not
hold good. Hence, jet wake flow pattern shown in Fig. 3
was not observed in the pumps (Tuzson 2000).
Effect of surface roughness
As stated earlier in this paper that the surface roughness
helps dropping performance and it is supposed to have less
performance if water is pumped. The following discussion
tries to highlight its effect for different viscosity fluid and
the design aspects of blade angles.
The properties the fluids (Table 2) were used to find the
effect of roughness and viscosity for different exit blade
angles on the pump performance. Characteristic curves for
Q = 0.1003 m3/s, N = 1470 rpm with ks = 0, 50 and
100 lm have been shown in Fig. 8. Increase in surface
roughness increases the head (H), input power (P) and
decreases the efficiency (g) (Varley 1961; Li 2008). It is
quite obvious that the pumping performance of the
roughened hub and shrouds is increased. This results in an
increase in head. A decrease in efficiency with higher ksvalues can be noticed. This occurs due to the flow losses in
the impeller passages and also due to external disc friction
(Varley 1961).
ω
Suction side Pressure side
Wakews
Jetw
Fig. 3 Jet-wake flow pattern in impeller
122 J Petrol Explor Prod Technol (2016) 6:117–127
123
Combined effect of exit blade angle and roughness
The results show that head generation increases due to the
combined effect of increase in b2 and ks values. Rough
surface implies a deceleration of relative velocity near the
solid walls with boundary layer thickening (Fig. 9). A large
b2 results in an increase in slip factor (Eq. 9). An increase inslip factor and head is caused by the lower relative velocity.
Also a roughness increase of the impeller passages result in a
small increase in head; in such cases the roughness has an
effect on the slip factor through an increase of the absolute
velocity in the boundary layers and an impact on secondary
flow. As the exit blade angle increases, the length of the
blades decreases resulting into greater friction losses in the
longer impeller passages. A large exit blade angle facilitates
the fluid to leave the impeller with a high absolute velocity
which is responsible for large mixing and great skin friction
Fig. 4 Performance curves for water at different blade angles (The
legend for each line/symbol shows that the first number is b1 and the
second number is b2.)
Fig. 5 Performance curves for different fluids
J Petrol Explor Prod Technol (2016) 6:117–127 123
123
losses. Hence a moderate fall in efficiency is accounted. The
combined effect of increase in b2 and ks values, the effi-
ciency gets affected. Because of increase in surface
roughness, the head loss (DH) is more due to the secondary
losses, profile losses and shock losses (Srinivasan 2008)
(Eq. 11). Equation 15 illustrates that the surface roughness
affects the friction and hence the impeller performance
drops. In addition, the decrease in efficiency to pump crude
oil and saline-water is due to the disc friction losses over the
Fig. 6 Static pressure contours at 50 % (Q = 0.1003 m3/s)
(a) Normal water
(b) Saline-water
(c) Crude oil
(d) Gasoline
(e) Kerosene
β2=25o β2=40o β2=70o
PSSS
Fig. 7 Flow velocity contours at 50 % span (Q = 0.1003 m3/s)
124 J Petrol Explor Prod Technol (2016) 6:117–127
123
outsides of the impeller shroud and hub. These results show a
good agreement with the existing analytical results (Gulich
2010; Zhang 2011).
In a total combined increase in b2 and ks values
increases the head and this leads to an increase in input
power with a moderate fall in efficiency.
Hence, it is recommended that in case of crude oil
pumping if the priority is to transport oil to a greater height,
high surface roughness value and large exit angle can be
introduced with a little compromise in efficiency. In other
scenarios, if power saving and a fair overall pumping
performance is a criterion, it is advised to select a moderate
roughness value and exit blade angle.
Conclusions
Effects of exit blade angle and surface roughness on the
centrifugal pump impeller performance at different mass
flow rates have been evaluated by numerical simulations
Hydraulic smooth Roughness= 50 µm Roughness= 100 µm
LE
TE
(a) Normal water
(b) Saline water
(c) Crude oil
(d) Kerosene
(e) Gasolene
Fig. 9 Meridional velocity contours for different fluids at design
point
Fig. 8 Effect of surface roughness (Q = 0. 1003 m3/s)
J Petrol Explor Prod Technol (2016) 6:117–127 125
123
for water, crude oil, saline-water gasoline and kerosene.
The conclusions are:
• The blade exit angle has higher influence on the head,
shaft power and hydraulic efficiency while the inlet
blade angle has lower effect on the parameters. Large
exit blade angle always augments head generation with
increased input power consumption. Increase in exit
blade angle increases the hydraulic efficiency till the
design point. At off-design points, the efficiency
decreases for higher exit blade angles because of
different losses.
• Higher viscosity liquids have lower head generation at
the same exit blade angle. An increase in fluid density
results in increased power consumption at various blade
exit angles.
• The efficiency drops with increase in surface roughness
due to surface roughness effect, flow losses and
external disc friction. Higher surface roughness helps
getting higher head.
• Combined effect of increase in exit blade angle and
surface roughness shows an in increase in head with
negligible increase in efficiency.
• In crude oil industry, for high head oil transportation
and more production, increase in surface roughness
value with large exit blade angle is effective. Hence,
prior to application, the design may consider design
itself if there is possibility of surface deterioration with
course of time. This will help reducing the total energy
consumption.
• From energy saving point of view and a fair overall
crude oil pumping performance, a moderate selection in
surface roughness value and exit blade angle is
suggested, and this can be implemented while design-
ing the component.
Acknowledgments The authors would like to acknowledge Indian
Institute of Technology Madras for the NFSC grant (Grant code:
OEC/10-11/529/NFSC/ABDU) to conduct this research.
Open Access This article is distributed under the terms of the
Creative Commons Attribution 4.0 International License (http://cre-
ativecommons.org/licenses/by/4.0/), which permits unrestricted use,
distribution, and reproduction in any medium, provided you give
appropriate credit to the original author(s) and the source, provide a
link to the Creative Commons license, and indicate if changes were
made.
References
Ansys-CFX 13.0 (2010) Reference Guide, Ansys Inc
Aoki K, Ohta H, Nakayama Y (1987) Study on centrifugal pump for
high viscosity liquids (The 1st Report, Effect of impeller output
angle and number of blades on the pump performance of closed
type impellers). Proceedings of the School of Engineering, Tokai
university 27(2):151–158
Bayly BJ, Orszag SA (1988) Instability mechanisms in shear flow
transition. Annu Rev Fluid Mech 20(1):359–391
Bellary SAI, Samad A (2014) An alternative approach to surrogate
averaging for a centrifugal impeller shape optimisation. Com-
puter Aided Engineering and Technology (Accepted for publi-
cation), Int. J
Bellary SAI, Husain A, Samad A (2014) Effectiveness of meta-
models for multi-objective optimization of centrifugal impeller.
J Mech Sci Technol 28(12):4947–4957
Church AH (1972) Centrifugal pump and blowers. 1st edn, John
Wiley and sons, Inc., New York
Dean RC Jr., Senoo Y (1960) Rotating wakes in vaneless
diffusers. ASME Journal of Basic Engineering September,
pp 563–574
Desai J, Chauhan V, Charnia S, Patel K (2011) Validation of
hydraulic design of a metallic volute centrifugal pump using
CFD. The 11th Asian International Conference on Fluid
Machinery and 3rd Fluid Power Technology Exhibition, 21–23
Nov., IIT Madras, Chennai
Fard MHS, Boyaghchi FA (2007) Studies on the influence of various
blade outlet angles in a centrifugal pump when handling viscous
fluids. Am J Appl Sci 4(9):718–724
Fard MHS, Tahani M, Ehghaghi MB, Fallahian MA, Beglari M
(2012) Numerical study of the effects of some geometric
characteristics of a centrifugal pump impeller that pumps a
viscous fluid. Comput Fluids 60:61–70
Gulich JF (2010) Centrifugal Pumps. Springer Publications, 2nd edn,
Berlin
Ippen AT (1946) The influence of viscosity on centrifugal perfor-
mance. Trans ASME 68(8):823–830
Itaya S, Nishikawa T (1960) Studies on the volute pumps handling
viscous fluids. Bull JSME 26(162):202–212
Johnston JP, Rothe PH (1967) Effect of system rotation on the
performance of two-dimensional diffusers. ASME J Fluids Eng,
pp 422–430
Johnston PJ, Halleen RM, Lezius DK (1972) Effects of spanwise
rotation on the structure of two-dimensional fully developed
channel flow. J Fluid Mech 56(3):533–557
Kamimoto G, Matsuoka Y (1956) On the flow in the impeller of
centrifugal type hydraulic machinery (The 2nd report). Trans
JSME Series 3, 22(113):55–59
Lazarkiewicz S, Troskolanski AT (1965) Impeller pumps. 1st edn,
Pergamon Press Ltd., Oxford
Li WG (2008) Numerical study on behavior of a centrifugal pump
when delivering viscous oils-part1: performance. Int J Turbo Jet
Engines 25:61–79
Li WG (2011a) Blade exit angle effects on performance of a
standard industrial centrifugal oil pump. J Appl Fluid Mech
4(2):105–119
Li WG (2011b) Effect of exit blade angle, viscosity and roughness in
centrifugal pumps investigated by cfd computations. Task
Quarterly 15(1):21–41
Luo X, Zhang Y, Peng L, Xu H, Yu W (2008) Impeller inlet geometry
effect on performance improvement for centrifugal pumps.
J Mech Sci Technol 22:1971–1976
Murakami M, Kikuyama K, Asakura E (1980) Velocity and pressure
distributions in the impeller passages of centrifugal pump.
ASME J. Fluids Eng 102:420–426
Ohta H, Aoki K (1996) Effect of impeller angle on performance and
internal flow of centrifugal pump for high viscosity liquids. Proc
School Eng Tokai Univ 36(1):159–168
Samad A, Kim KY (2008) Shape optimization of an axial compressor
blade by multi-objective genetic algorithm. Proc. IMechE Part
A. J. Power Energy 222:599–611
126 J Petrol Explor Prod Technol (2016) 6:117–127
123
Sanda B, Daniela CV (2012) The influence of the inlet angle over the
radial impeller geometry design approach with Ansys. J Eng
Studies Res 18(4):32–39
Shigemitsu T, Fukutomi J, Nasada R, Kaji K (2011) The effect of
blade outlet angle on performance and internal flow condition of
mini turbo-pump. J Therm Sci 20(1):32–38
Srinivasan KM (2008) Rotodynamic pumps. New Age International
(P) Ltd., New Delhi
Stepanoff AJ (1940) Pumping viscous oils with centrifugal pumps.
Oil Gas J 1(4)
Stepanoff AJ (1964) Centrifugal and axial flow pumps. 2nd edn, John
Wiley and Sons inc., New York
Telow N (1942) A survey of modern centrifugal pump practice for
oilfield and Oil Refinery Services. The Institution of Mechanical
Engineering 3(121)
Tuzson J (1993) Interpretation of impeller flow calculations. ASME J
Fluid Eng, pp 463–467
Tuzson J (2000) Centrifugal pump design. 1st edn, John Wiley and
Sons, Inc., New York
Varley FA (1961) Effects of impeller design and surface roughness on
the performance of centrifugal pumps. Proc Instn Mech Eng
175(21):955–969
Wilcox DC (1994) Turbulence modeling for CFD. 2nd edn, DCW
Industries, Inc., La Canada, California
Zhang Y, Zhou X, Ji Z, Jiang C (2011) Numerical design and
performance prediction of low specific speed centrifugal pump
impeller. Int J Fluid Machin Syst 4(1):133–139
J Petrol Explor Prod Technol (2016) 6:117–127 127
123